The present invention relates to the automotive timing chain art. It finds particular application in conjunction with a unidirectional roller chain sprocket for use in automotive camshaft drive applications and will be described with particular reference thereto. However, the present invention may also find application in conjunction with other types of chain drive systems and applications where reducing the noise levels associated with chain drives is desired.
Roller chain sprockets for use in camshaft drives of automotive engines are typically manufactured according to ISO (International Organization for Standardization) standard 606:1994(E). The ISO-606 standard specifies requirements for short-pitch precision roller chains and associated chain wheels or sprockets.
FIG. 1 illustrates a symmetrical tooth space form or an ISO-606 compliant sprocket. The tooth space has a continuous fillet or root radius R.sub.i extending from one tooth flank (i.e., side) to the adjacent tooth flank as defined by the roller seating angle .alpha.. The flank radius R.sub.f is tangent to the roller seating radius R.sub.1 at the tangency point TP. A chain with a link pitch P has rollers of diameter D.sub.1 in contact with the tooth spaces. The ISO sprocket has a chordal pitch also of length P, a root diameter D.sub.2, and Z number of teeth. The pitch circle diameter PD, tip or outside diameter OD, and tooth angle A (equal to 360.degree./Z) further define the ISO-606 compliant sprocket. The maximum and minimum roller seating angle .alpha. is defined as: EQU .alpha..sub.max =(140.degree.-90.degree.)/Z and .alpha..sub.min =(120.degree.-90.degree.)/Z
With reference to FIG. 2, an exemplary ISO-606 compliant roller chain drive system 10 rotates in a clockwise direction as shown by arrow 11. The chain drive system 10 includes a drive sprocket 12, a driven sprocket 14 and a roller chain 16 having a number of rollers 18. The sprockets 12, 14, and chain 16 each generally comply with the ISO-606 standard.
The roller chain 16 engages and wraps about sprockets 12 and 14 and has two spans extending between the sprockets, slack strand 20 and taut strand 22. The roller chain 16 is under tension as shown by arrows 24. A central portion of the taut strand 22 may be guided between the driven sprocket 14 and the drive sprocket 12 with a chain guide 26. A first roller 28 is shown at the onset of meshing at a 12 o'clock position on the drive sprocket 12. A second roller 30 is adjacent to the first roller 28 and is the next roller to mesh with the drive sprocket 12.
Chain drive systems have several components of undesirable noise. A major source of roller chain drive noise is the sound generated as a roller leaves the span and collides with the sprocket during meshing. The resultant impact noise is repeated with a frequency generally equal to that of the frequency of the chain meshing with the sprocket. The loudness of the impact noise is a function of the impact energy (E.sub.A) occurring during the meshing process. The impact energy (E.sub.A) is related to engine speed, chain mass, and the impact velocity between the chain and the sprocket at the onset of meshing. The impact velocity is affected by the chain-sprocket engagement geometry, of which an engaging flank pressure angle .gamma. (FIG. 3) is a factor, where: ##EQU1## E.sub.A =Impact Energy [N.multidot.m]V.sub.A =Roller Impact Velocity [m/s]
.gamma.=Engaging Flank Pressure Angle PA1 n=Engine Speed [RPM] PA1 w=Chain Mass [Kg/m] PA1 Z=Number of Sprocket Teeth PA1 A=Tooth Angle (360.degree./Z) PA1 .alpha.=Roller Seating Angle PA1 P=Chain Pitch (Chordal Pitch)
The impact energy (E.sub.A) equation presumes the chain drive kinematics will conform generally to a quasi-static analytical model and that the roller-sprocket driving contact will occur at a tangent point TP (FIG. 3) of the flank and root radii as the sprocket collects a roller from the span.
As shown in FIG. 3, the pressure angle .gamma. is defined as the angle between a line A extending from the center of the engaging roller 28, when it is contacting the engaging tooth flank at the tangency point TP, through the center of the flank radius R.sub.f, and a line B connecting the centers of the fully seated roller 28, when it is seated on root diameter D.sub.2 (Point A), and the center of the next meshing roller 30, as if it were also seated on root diameter D.sub.2 in its engaging tooth space. The roller seating angles .alpha. and pressure angles .gamma. listed in FIG. 25 are calculated from the equations defined above. It should be appreciated that .gamma. is a minimum when .alpha. is a maximum. The exemplary 18-tooth, ISO-606 compliant, sprocket 12 of FIG. 3 will have a pressure angle .gamma. in the range of 12.5.degree. to 22.5.degree. as listed in the table of FIG. 25.
FIG. 3 also shows the engagement path (phantom rollers) and the driving contact position of roller 28 (solid) as the drive sprocket 12 rotates in the direction of arrow 11. FIG. 3 depicts the theoretical case with chain roller 27 seated on root diameter D.sub.2 of a maximum material sprocket with both chain pitch and sprocket chordal pitch equal to theoretical pitch P. For this theoretical case, the noise occurring at the onset of roller engagement has a radial component F.sub.R as a result of roller 28 colliding with the root surface R.sub.i and a tangential component F.sub.T generated as the same roller 28 collides with the engaging tooth flank at point TP as the roller moves into driving contact. It is believed that the radial impact occurs first, with the tangential impact following nearly simultaneously. Roller impact velocity V.sub.A is shown to act through, and is substantially normal to, engaging flank tangent point TP with roller 28 in driving contact at point TP.
The impact energy (E.sub.A) equation accounts only for a tangential roller impact during meshing. The actual roller engagement, presumed to have a tangential and radial impact (occurring in any order), would therefore seem to be at variance with the impact energy (E.sub.A) equation. The application of this quasi-static model, which is beneficially used as a directional tool, permits an analysis of those features that may be modified to reduce the impact energy occurring during the tangential roller-sprocket collision at the onset of meshing. The radial collision during meshing, and its effect on noise levels, can be evaluated apart from the impact energy (E.sub.A) equation.
Under actual conditions as a result of feature dimensional tolerances, there will normally be a pitch mismatch between the chain and sprocket, with increased mismatch as the components wear in use. This pitch mismatch serves to move the point of meshing impact, with the radial collision still occurring at the root surface R.sub.i but not necessarily at D.sub.2. The tangential collision will normally be in the proximity of point TP, but this contact could take place high up on the engaging side of root radius R.sub.i or even radially outward from point TP on the engaging flank radius R.sub.f as a function of the actual chain-sprocket pitch mismatch.
Reducing the engaging flank pressure angle .gamma. reduces the meshing noise levels associated with roller chain drives, as predicted by the impact energy (E.sub.A) equation set forth above. It is feasible but not recommended to reduce the pressure angle .gamma. while maintaining a symmetrical tooth profile, which could be accomplished by simply increasing the roller seating angle .alpha., effectively decreasing the pressure angle for both flanks. This profile as described requires that a worn chain would, as the roller travels around a sprocket wrap (discussed below), interface with a much steeper incline and the rollers would necessarily ride higher up on the coast flank prior to leaving the wrap.
Another source of chain drive noise is the broadband mechanical noise generated in part by shaft torsional vibrations and slight dimensional inaccuracies between the chain and the sprockets. Contributing to a greater extent to the broadband mechanical noise level is the intermittent or vibrating contact that occurs between the unloaded rollers and the sprocket teeth as the rollers travel around the sprocket wrap. In particular, ordinary chain drive system wear comprises sprocket tooth face wear and chain wear. The chain wear is caused by bearing wear in the chain joints and can be characterized as pitch elongation. It is believed that a worn chain meshing with an ISO standard sprocket will have only one roller in driving contact and loaded at a maximum loading condition.
With reference again to FIG. 2, driving contact at maximum loading occurs as a roller enters a drive sprocket wrap 32 at engagement. Engaging roller 28 is shown in driving contact and loaded at a maximum loading condition. The loading on roller 28 is primarily meshing impact loading and the chain tension loading. The next several rollers in the wrap 32 forward of roller 28 share in the chain tension loading, but at a progressively decreasing rate. The loading of roller 28 (and to a lesser extent for the next several rollers in the wrap) serves to maintain the roller in solid or hard contact with the sprocket root surface 34.
A roller 36 is the last roller in the drive sprocket wrap 32 prior to entering the slack strand 20. Roller 36 is also in hard contact with drive sprocket 12, but at some point higher up (e.g., radially outwardly) on the root surface 34. With the exception of rollers 28 and 36, and the several rollers forward of roller 28 that share the chain tension loading, the remaining rollers in the drive sprocket wrap 32 are not in hard contact with the sprocket root surface 34, and are therefore free to vibrate against the sprocket root surfaces as they travel around the wrap, thereby contributing to the generation of unwanted broadband mechanical noise.
A roller 38 is the last roller in a sprocket wrap 40 of the driven sprocket 14 before entering the taut strand 22. The roller 38 is in driving contact with the sprocket 14. As with the roller 36 in the drive sprocket wrap 32, a roller 42 in the sprocket wrap 40 is in hard contact with a root radius 44 of driven sprocket 14, but generally not at the root diameter.
It is known that providing pitch line clearance (PLC) between sprocket teeth promotes hard contact between the chain rollers and sprocket in the sprocket wrap, even as the roller chain wears. The amount of pitch line clearance added to the tooth space defines a length of a short arc that is centered in the tooth space and forms a segment of the root diameter D.sub.2. The root fillet radius R.sub.i is tangent to the flank radius R.sub.F and the root diameter arc segment. The tooth profile is still symmetrical, but R.sub.i is no longer a continuous fillet radius from one flank radius to the adjacent flank radius. This has the effect of reducing the broadband mechanical noise component of a chain drive system. However, adding pitch line clearance between sprocket teeth does not reduce chain drive noise caused by the roller-sprocket collision at impact.
Chordal action, or chordal rise and fall, is another important factor affecting the operating smoothness and noise levels of a chain drive, particularly at high speeds. Chordal action occurs as the chain enters the sprocket from the free span during meshing and it can cause a movement of the free chain in a direction perpendicular to the chain travel but in the same plane as the chain and sprockets. This chain motion resulting from chordal action will contribute an objectionable noise level component to the meshing noise levels, so it is therefore beneficial to reduce chordal action inherent in a roller chain drive.
FIGS. 4a and 4b illustrate the chordal action for an 18-tooth, ISO-606 compliant, sprocket having a chordal pitch of 9.525 mm. Chordal rise 45 may conventionally be defined as the displacement of the chain centerline as the sprocket rotates through an angle A/2, where: EQU Chordal rise=r.sub.p =r.sub.c =r.sub.p [1-cos(180.degree./Z)]
where r.sub.c is the chordal radius, or the distance from the sprocket center to a pitch chord of length P; r.sub.p is the actual theoretical pitch radius; and Z is the number of sprocket teeth.
It is known that a short pitch chain provides reduced chordal action compared to a longer pitch chain having a similar pitch radius. FIGS. 4a and 4b show only the drive sprocket and assume a driven sprocket (not shown) also having 18-teeth and in phase with the drive sprocket shown. In other words, at T=0 (FIG. 4a), both sprockets will have a tooth center at the 12 o'clock position. Accordingly, this chain drive arrangement under quasi-static conditions will have a top or taut strand that will move up and down in a uniform manner a distance equal to that of the chordal rise. At T=0, a roller 46 is at the onset of meshing, with chordal pitch P horizontal and in line with taut strand 22. At T=0+(A/2), (FIG. 4b), roller 46 has moved to the 12 o'clock position.
For many chain drives, the drive and driven sprockets will be of different sizes and will not necessarily be in phase. The chain guide 26 (FIG. 2) has the primary purpose to control chain strand vibration in the taut span. The geometry of the guide-chain interface also defines the length of free span chain over which chordal rise and fall is allowed to operate. FIG. 5 is an enlarged view of FIG. 2 showing the first roller 28 at the onset of engagement and the second roller 30 as the next roller about to mesh with sprocket 12. In this example, the chain guide 26 controls and guides the engaging portion of the taut strand 22 except for five (5) unsupported or "free" link pitches extending between the chain guide 26 and the engaging roller 28. This length of unsupported link pitches for the engaging portion of taut strand 22 in this example is horizontal when roller 28 is at the 12 o'clock position.
With reference to FIG. 6, the drive sprocket 12 is rotated in a clockwise direction to advance roller 28 to a new angular position (A/2)+.omega., where .omega. is the added rotation angle as determined by a quasi-static engagement geometry with roller 28 being fully seated and roller 30 is at the instant of sprocket engagement. The roller 28 is considered to be seated and in hard contact with the root surface at D.sub.2 at the onset of meshing of roller 30, and a straight line is assumed for the chain span from roller 28 to a chain pin center 48, about which the unsupported or "free" span from pin 48 to engaging roller 30 is considered to rotate.
As a result of the chordal action, the engaging free span is no longer horizontal to satisfy the roller engaging geometry. This is in contrast to the chain drive as described in FIG. 4a in which chordal action causes the taut strand to move uniformly, but in a horizontal path because the drive and driven sprockets have the same number of teeth and the sprocket teeth are in phase. It should be appreciated that the straight line assumption is valid only in a quasi-static model. The amount of movement or deviation from the straight line assumption will be a function of the drive dynamics, the chain control devices, and the chain drive and sprocket geometry. The location and chain-interfacing contour of the chain guide 26 will determine the number of free span pitches for which chain motion will take place as a result of the chordal rise and fall during the roller meshing process.
Assuming that rollers 28 and 30 are in hard contact with the sprocket root surfaces at 50 and 52, respectively, the chordal rise is the perpendicular displacement of the center of roller 30 (located on the pitch diameter PD) from the taut span 22 path as it moves from its initial meshing position shown to the 12 o'clock position. For the chain drive arrangement as described, there is no modulation to the roller meshing frequency. In other words, roller meshing for all sprocket teeth will occur at substantial y the same frequency for a constant engine speed.
Accordingly, it is desirable to develop a new and improved roller chain drive system which meets the above-stated needs and overcomes the foregoing disadvantages and others while providing better and more advantageous results.